Hydraulic brake pressure regulator

ABSTRACT

In the particular embodiments of the invention described in the specification, a hydraulic brake pressure regulator for the hydraulic brake system of a motor vehicle is designed to provide a family of characteristic curves which are dependent upon the vehicle load (outlet pressure as a function of inlet pressure), in which each characteristic curve follows a nonlinear and constant course, so as to provide optimum conformance with the ideal braking-force distribution of the vehicle.

BACKGROUND OF THE INVENTION

This invention relates to hydraulic brake pressure regulators forhydraulic brake systems of motor vehicles, and, more particularly, tobrake pressure regulators having a family of characteristic curvesdependent upon a vehicular parameter such as outlet pressure as afunction of inlet pressure.

Such brake pressure regulators are used to obtain braking performancefor a motor vehicle which is independent of variable vehicularparameters, such as the load on the rear axle or the deceleration of thevehicle, to provide as nearly as possible a constant braking performancefor the vehicle.

It is common knowledge that when a motor vehicle is braked the frontaxle becomes more heavily loaded, in comparison with its static load,and the rear axle is correspondingly unloaded. As is known, optimumbraking of a vehicle is obtained when the front and rear axles requirethe same frictional brake contact at particular road conditions and at aparticular speed. This is the case when, for instance, the front andrear axles, on braking, simultaneously reach the transition region forlocking.

When the dimensions, weight and location of the center of gravity of thevehicle are known, the optimum distribution of the braking force betweenthe front and rear axles may be calculated. This ideal braking-force orbraking-pressure distribution represented by P₁ (front) and P₂ (rear),referred to the weight G of the vehicle, is shown qualitatively in FIG.1 for the "full load" condition by the upper thin solid line and for the"empty load+driver" condition by the lower solid line. It may be seenthat continuous nonlinear characteristic curves are involved here. Aseparate curve (not shown), lying between these upper and lower boundarylines, applies to each intermediate load condition of the vehicle.

Locking of the rear wheels before the front wheels are locked should beavoided in every load condition because it results in unstable drivingperformance. To avoid such overbraking of the rear axle, thedistribution of braking forces to the front and rear axles must beselected in designing a hydraulic brake system so that the effectivebraking force on the rear axle is never higher than the braking forceshown in FIG. 1 for the case of ideal braking.

It would thus be possible to avoid overbraking of the rear axle insimple fashion by providing a fixed braking-force distribution whichcould, for example, be aimed at making the real braking-forcedistribution at braking, a=0.8, a point of the ideal braking-forcedistribution for the load case "empty load+driver" i.e., the lower thinline shown in FIG. 1. In such a fixed braking-force distribution, thestraight line represented in FIG. 1 as a heavy solid line would then beproduced.

It is easy to see that, with such a fixed braking-force distribution,the ideal conditions would be present only for this special load caseand for the assumed braking a=0.8, whereas with greater loading of thevehicle or with a smaller demand for frictional contact conditions farfrom ideal braking performance would result.

In order to at least approximate the actual braking-force distributionof a motor vehicle to the ideal braking-force distribution, pressureregulators having a bent characteristic line with a knee position whichvaries according to the load, are disclosed, for example in DE-AS No.1,655,003, FIG. 3, and are already included in many vehicles. Typicalcharacteristic bent lines of such a known brake pressure regulator areindicated by dashed lines in FIG. 1. Even when such a knownload-dependent brake pressure regulator is used, the real braking-forcedistribution is still clearly far from the ideal distribution.

A better approximation of the ideal braking-force distribution isprovided by hydraulic brake regulators whose family of characteristiclines consists of a group of bent straight lines in which both the slopeand the knee are variable in accordance with the load. Such brakepressure regulators are disclosed in, for example, DE-OS No. 2,708,941and DE-OS No. 2,923,018. The two boundary lines of such known brakepressure regulators, which might alternatively be called characteristicline regulators having a load-dependent knee, are indicated bydash-dotted lines in FIG. 1. But even these known hydraulic brakepressure regulators having their degressive characteristic lines stillleave a great deal to be desired with respect to their adaptability toideal braking-force distribution. In addition, they are comparativelycomplicated in design. One known brake pressure regulator of this type(DE-OS No. 2,923,018) works, for example, like known characteristicregulators (DE-OS No. 1,780,560) using the principle of the scalebalance, in which the effective lever lengths of a scale balancecooperating with two piston members are automatically varied as afunction of the vehicle load. This brake pressure regulator is not onlymechanically complicated, but likewise has a comparatively large size.

The object of the invention is to provide an improved hydraulic brakepressure regulator for the hydraulic brake system of a motor vehicle,having a family of characteristic curves dependent upon a vehicleparameter such as outlet pressure as a function of inlet pressure.

Another object of the invention is to provide a hydraulic pressureregulator which not only provides a family of characteristic curvesadapted as closely as possible to the ideal braking-force distributionbut also has comparatively low construction and manufacturing costs andsmall spatial requirements.

SUMMARY OF THE INVENTION

These and other objects of the invention are attained by providing ahydraulic brake pressure regulator for the hydraulic brake system of amotor vehicle especially designed to provide a family of characteristiccurves which are dependent upon a vehicle parameter such as outletpressure as a function of inlet pressure in which each curve of thefamily is a continuous non-linear curve which may be degressive orprogressive. A particular form of regulator prepared in accordance withthe invention includes a first chamber with an inlet and a secondchamber with an outlet separated by an axially displaceable first pistonhaving a control valve which may be opened and closed to permit aconnection between the first and second chambers and including a secondpiston displaceable independently of the first piston to define thesecond chamber, the second piston being coupled to the control valve sothat the opening and closing of the control valve is dependent upon thespacing of the two pistons, one piston also being coupled to anonhydraulic force mechanism having a linear characteristic and theother pistons being coupled to a nonhydraulic second force mechanismwhich is dependent upon a vehicle parameter. One of the force mechanismsacts on the first piston in a direction opposing pressure appliedthrough the first inlet and the other acts on the second piston in thesame direction. The brake regulator of the invention may be inserted inthe hydraulic pressure line between the main brake cylinder and thebrake cylinder of the rear axle or between the main brake cylinder andthe brake cylinder of the front axle of the vehicle and the first forcemechanism may be a linear spring while the second force mechanism may bea pneumatic piston.

BRIEF DESCRIPTION OF THE DRAWINGS

The invention and preferred embodiments thereof are explainedhereinafter in greater detail with reference to the accompanyingdrawings, in which:

FIG. 1 is a graphical representation showing a family of characteristiclines of the braking-force distribution of a motor vehicle;

FIG. 2 is a view in longitudinal section illustrating a first embodimentof a brake pressure regulator pursuant to the invention forincorporation into the brake circuit of the front axle of a vehicle;

FIG. 3 is a view in longitudinal section, similar to that of FIG. 2,illustrating a second embodiment of a brake pressure regulator accordingto the invention;

FIG. 4 is a graphical representation showing a family of characteristiccurves, obtainable with the brake pressure regulators of FIGS. 2 and 3,illustrating the relation between the regulator outlet pressure and theregulator inlet pressure related to the weight of the vehicle;

FIG. 5 is a view in longitudinal section illustrating a first embodimentof a brake pressure regulator for incorporation into the brake circuitof the rear axle of a vehicle;

FIG. 6 is a view in longitudinal section similar to that of FIG. 5showing another embodiment of the invention;

FIG. 7 is a view in longitudinal section similar to that of FIG. 5showing a further embodiment of the invention;

FIG. 8 is a graphical representation showing a family of characteristiccurves, obtainable with the brake pressure regulators of FIGS. 5-7,illustrating the relation between the regulator outlet pressure and theregulator inlet pressure related to the weight of the vehicle;

FIG. 9 is a schematic diagram illustrating an example of a hydraulicbrake system with a brake pressure regulator pursuant to FIG. 2 or FIG.3, arranged in a brake circuit for the front axle of a vehicle; and

FIG. 10 is a schematic diagram showing an embodiment of a hydraulicbrake system with a hydraulic brake pressure regulator pursuant to anyof FIGS. 5-7, arranged in the brake circuit for the rear axle of avehicle.

DESCRIPTION OF PREFERRED EMBODIMENTS

A typical hydraulic brake pressure regulator pursuant to the inventionas shown in FIG. 2 consists of a regulator housing 1 with a firstpressure chamber 2 and a second pressure chamber 4, which are separatedfrom each other by an axially displaceable first piston 6. The firstpressure chamber 2 has a pressure inlet 3 through which a hydraulicmedium having an inlet pressure P₁ supplied, for example, from a mainbrake cylinder of the motor vehicle, is introduced into the firstpressure chamber. The second pressure chamber 4 has a pressure outlet 5through which the hydraulic fluid may be supplied, at an outlet pressureP₂, which is lower than the inlet pressure P₁, to hydraulically operateddevices such as the front or rear wheel brake cylinders of the brakesystem.

Spaced from the first piston 6 is a second piston 8, which is axiallymovable independently of the first piston and which limits the secondpressure chamber 4 in the axial direction. A regulating valve 7, havinga valve member 71, cooperating with a valve bore 72 is arranged in thefirst piston 6, and a rod-shaped force transmitting member 9 extendsfrom the second piston 8 to releasably engage the valve member 71. Thusthe opening and closing of the control valve 7 depends upon the axialdistance between the two pistons 6 and 8. Opening and closing of thecontrol valve 7 intermittently connects the chambers 2 and 4, permittinghydraulic fluid to flow between them and transmitting pressure from onechamber to the other.

Each of the two pistons 6 and 8 is acted upon by a non-hydraulicopposing force, the two forces being independent of one another. Theseopposing forces are directed so that they counteract the force exertedby the inlet pressure P₁ on the first piston 6 and/or the force exertedby the outlet pressure P₂ on the second piston 8. One opposing force isproduced by a nonhydraulic first force mechanism in the form of amechanical elastic device such as a spring 10 having a linearcharacteristic and the second opposing force is produced by anotherhydraulic force mechanism, independent of the device 10, in the form ofa pneumatic elastic device 11 having a progressive characteristic. Whilethe progressive characteristic of the second force mechanism, that is,the pneumatic elastic device 11, is variable as a function of vehicleparameter, in particular, as a function of vehicle load, the linearcharacteristic of the first elastic mechanism, the mechanical elasticdevice 10, is invariable.

A hydraulic brake pressure regulator designed in this fashion has afamily of characteristic curves which are dependent upon a vehicleparameter i.e., the outlet pressure of the regulator as a function ofinlet pressure, the individual curves of which follow a nonlinear andentirely continuous curve i.e., one with no knee, so that it becomespossible to adapt the actual braking-force distribution of a vehicleequipped with such a hydraulic brake pressure regulator very closely tothe ideal braking-force distribution.

Depending on which of the two different nonhydraulic force mechanismsthe two pistons 6 and 8 cooperate with, the nonlinear curves of thefamily of characteristic curves (outlet pressure as a function of inletpressure) will follow a course which is either degressive, i.e.continuously decreasing with respect to a straight-line relation, orprogressive i.e., continuously increasing with respect to astraight-line relation. Thus, to obtain the desired almost idealbraking-force distribution, brake pressure regulators according to theinvention having a degressive course of characteristic curves may beconnected in the customary manner between the main brake cylinder andthe wheel brake cylinders of the rear axle of the vehicle, as indicatedschematically in FIG. 10.

On the other hand, hydraulic brake pressure regulators for obtaining thealmost ideal braking-force distribution which have a family ofprogressive characteristic lines must be connected between the mainbrake cylinder and the wheel brake cylinders of the front axle of thevehicle, as indicated schematically in FIG. 9. While the arrangement ofFIG. 10 may preferably be utilized in ordinary passenger cars, thearrangement according to FIG. 9 is preferably used in trucks, in whichthe difference between full load and empty load is especially great.FIGS. 2 and 3 show two representative brake pressure regulatorsaccording to the invention by which a family of characteristic curvesfollowing a progressive course may be obtained.

FIG. 4 is a graphical representation illustrating one such family ofprogressive characteristic curves where the outlet pressure P₂ and theinlet pressure P₁ are in each instance related to the weight G of thevehicle and only the two boundary curves "full load" and "emptyload+driver" are shown.

In the two brake pressure regulators according to FIGS. 2 and 3, thefirst force mechanism having a linear characteristic, i.e., themechanical elastic device 10, cooperates with the first piston 6 in FIG.2 or 6a in FIG. 3. In these examples, the device 10 is a compressionspring, having one end engaging a spring ring, and the other endengaging the head of the piston 6 or 6a.

The nonhydraulic second force mechanism having a progressivecharacteristic, namely the pneumatic elastic device 11, cooperates withthe second piston 8. It has a compression chamber filled with acompressible gaseous medium which is variable in size in accordance witha vehicle parameter, such as the vehicle load. In these embodiments thecompression chamber of the pneumatic elastic device 11 is designed as acylindrical chamber integrated into the regulator housing 1, with anaxially displaceable first cylinder end piston 12, the axial position ofwhich is displaceable in accordance with a parameter of the vehicle, forexample, the vehicle load, and with an axially displaceable secondcylinder end member which is connected to one of the two pistons. In theembodiments of FIGS. 2 and 3 this axially displaceable second cylinderend member of the compression chamber is formed by the second piston 8itself.

By controlling the size of the compression chamber of the pneumaticelastic device 11 the effective characteristic curve of elasticity ofthis elastic device may be selectively varied. In FIG. 2 thispossibility is indicated schematically by a rod 18, by means of whichthe first cylinder end 12 may be axially shifted further to the left orto the right, the type of adjusting force basically being of nosignificance to the invention. It might, for example, be producedmechanically or electrically, in particular, by a non-reactive drive byway of a spindle.

Instead of displacing the first cylinder and 12 of the compressionchamber mechanically or electromechanically, it may alternatively bedisplaced pneumatically or hydraulically, as shown, for example, in theembodiments of FIGS. 3 and 5 and 6. In these embodiments the brakepressure regulator has an additional pressure inlet, namely an adjustinginlet 19, 19' or 19" by way of which a pressure medium acting directlyor indirectly on the first cylinder end 12, 12' or 12" may be introducedto enlarge or reduce the compression chamber 11, 11' or 11" i.e., byaxial displacement of the cylinder end 12, 12' or 12".

In the embodiment according to FIG. 3, the pressure medium supplied byway of the adjusting inlet 19 acts by way of the piston rod 28 directlyon the first cylinder end 12, and specifically acts against the force ofa tension spring device 29 in the opposite direction on the cylinderend.

To give the brake pressure regulator a desired vehicle load-dependentfamily of characteristic curves, the size of the compression chamber 11,i.e., the distance designated "l₀ " in the drawings between the cylinderend 12 and the piston 8, must be varied according to the vehicle loadand, specifically, such that at the empty load+driver condition, l₀ hasa selected minimum value and at full load l₀ has a selected maximumvalue.

In the embodiment shown in the example of FIG. 3, a pressure medium isused for displacement of the axial position of the first cylinder end12. The pressure applied by the pressure medium to the face of the rod28 is inversely proportional to the weight of the vehicle, i.e., of theload of the vehicle, so that the displacing force acting on the face ofthe piston 28 is smaller at the full load condition than at the emptyload+driver condition.

The hydraulic brake pressure regulators illustrated in FIGS. 2 and 3 areessentially identical. Only to indicate various conceivable designpossibilities, the embodiment of FIG. 3, unlike that of FIG. 2, has itsfirst piston 6a constructed as a stepped piston and its first cylinderend 12 adjusted hydraulically, rather than mechanically, in its axialposition.

In operation of the embodiments shown in FIGS. 2 and 3, when the brakesystem is unpressured, i.e., the brakes are not actuated, the pistons 6,6a and 8, as well as the first cylinder end 12, are in positionsillustrated in the drawings. The mechanical elastic device 10 having alinear characteristic, designed as a compression spring, is slack. Thesecond piston 8 is axially positioned so that an equalizing channel 20,communicating with the compression chamber 11, is open. In theillustrated examples air is used as the compressible gaseous medium; itis therefore easily possible to lead the equalizing channel 20 to theoutside of the regulator so that the initial pressure P₀ prevailing inthe compression chamber 11 at this point of time corresponds to theatmospheric pressure. If for some reason another compressible gaseousmedium is used, the equalizing channel must lead, accordingly, to acorresponding gas reservoir with a defined initial pressure.

The axial position of the first cylinder end 12 is positioned accordingto the load of the vehicle, that is, for example, according to the levelof the vehicle with respect to the rear axle. Specifically, as alreadystated, the cylinder end 12 is positioned so that the compressionchamber 11 formed between the second piston 8 and the first cylinder end12 is large at high load and correspondingly small at low load. In thisconfiguration the valve member 71 of the control valve arranged in thefirst piston 6 lies on its valve seat from play, so that the valve bore72 is closed.

When the brake pedal is actuated, a pressure is developed in the mainbrake cylinder of the motor vehicle, which is supplied as the inletpressure P₁ at the pressure inlet of the brake pressure regulator. Thepressure P₁ acts upon the piston face F₁ of the first piston 6 whichfaces the first pressure chamber 2. As a result, the first piston ismoved a distance X toward the right as shown in the drawing. This motionof the first piston 6 causes the valve member 71 of the control valve,coupled by way of the rod-shaped force-transmitting member 9 with thesecond piston 8, to be raised from its valve seat, so that the hydraulicpressure medium may flow through the valve bore 72 into the secondpressure chamber 4, in which a pressure P₂ develops.

Thus, the pressure P₂ acts simultaneously on the first piston 6 and onthe second piston 8, so that, on the one hand, the first piston 6 ismoved a small distance opposite the X direction and, on the other, thesecond piston 8 is moved in the X direction. The valve member 71 of thecontrol valve then returns to its valve seat and, at the same time, theaperture of the equalizing channel 20, which opens into the compressionchamber 11, is shut off by the second piston 8, so that the compressionchamber 11 becomes effective as a pneumatic elastic mechanism having aprogressive characteristic.

The deflection of the two pistons 6 and 8, which are moved essentiallyindependently of one another, continues until the forces applied to themare in equilibrium, after which the control valve 7 is closed and thetwo pistons have travelled equidistant paths. This state of equilibriumis determined by the effective pressure and the elastic forces appliedto the two pistons 6 and 8.

In the stationary state of equilibrium the following equations apply:

    P.sub.1 ·F.sub.1 -P.sub.2 ·F.sub.2 -k·X=0 (1)

and:

    P.sub.2 ·F.sub.3 -P.sub.i ·F.sub.4 =0    (2)

where:

P₁ =inlet pressure of the brake pressure regulator

P₂ =outlet pressure of the brake pressure regulator

P_(i) =pressure in the compression chamber 11 on compression of itscompressible gaseous medium

F₁ =effective piston area of the first piston 6 facing the firstpressure chamber 2

F₂ =effective piston area of the first piston 6 facing the secondpressure chamber 4

F₃ =effective piston area of the second piston 8 facing the secondpressure chamber

F₄ =effective piston area of the second piston 8 at the second cylinderend

k=force constant of the mechanical elastic device 10

X=displacement of the first piston 6.

As soon as the aperture of the equalizing channel 20 opening into thecompression chamber 11 is closed off by the second piston 8, the gaseousmedium found in the said chamber is compressed by further motion of thesecond piston 8 in the X direction. If we assume a polytropic change ofstate for this compression, then the compression is effected accordingto the equation: ##EQU1## from which it follows that ##EQU2## where,inter alia: P₀ =initial pressure, specifically atmospheric pressure, inthe compression chamber 11 before compression

V₀ =initial volume of compression chamber

V_(i) =volume of the compression chamber 11 after compression

l₀ =initial axial length of the compression chamber (distance of thesecond piston 8 from the first cylinder end 12)

n=polytropic exponent.

Finally, substituting the expression for P_(i) from Equation 3 inEquation 2 and substituting the expression for X derived from Equation 2in Equation 1, the following expression is obtained: ##EQU3##

With l₀ as a variable parameter, a family of characteristic curves, asillustrated by the envelope in FIG. 4, for the brake pressure regulatorrepresented in FIGS. 2 and 3 follows from the above equation. Usingappropriate sizes for the various piston areas F₁ to F₄ (including, forexample, a stepped piston and stepped cylinder arrangement), andappropriate force constants k for the mechanical elastic device 10 and,if desired, an initial system pressure P₀ >P_(atmospheric), possibly inconjunction with an initial tension in the spring, all possiblecharacteristic curves may be obtained with the curvature tendency shownin FIG. 4 using all possible quasilinear characteristic lines with l₀ asa variational parameter. The characteristic of the regulator is, ofcourse, also influenced by the polytropic exponent n, i.e., by the typeof compressible gaseous medium used.

Because of the progressive characteristic of the pneumatic elasticmechanism formed by the compression chamber 11, the position of thesecond piston 8 does not change linearly with the outlet pressure P₂.The result of this, in turn, is that the outlet pressure P₂ is anonlinear function of the inlet pressure P₁. The ratio between thelinearly increasing spring force k·X of the mechanical elastic mechanism10, which acts on the first piston 6 or 6a, and the progressivelyincreasing compressive force P_(i) ·F₄ of the pneumatic elasticmechanism 11 varies according to the magnitude of the initial volume V₀or the initial axial length l₀ of the compression chamber 11, because ofthe fixed path connection between the first piston 6 or 6a and thesecond piston 8. Consequently l₀ is the determining value for thecorresponding initial rise of the outlet pressure P₂ in the family ofcharacteristic lines of FIG. 4. When the initial length l₀ or theinitial volume V₀ of compression chamber 11 is varied as a function ofvehicle load, as previously explained, a family of characteristic curvesis produced having constant and nonlinear characteristic courses withthe vehicle load as the variable parameter. Since the course, i.e., thecurvature, of each of the characteristic curves may be controlled by thevalues used for the various regulator pistons and elastic mechanisms,the family of characteristic curves of the brake pressure regulators maybe very closely approximated to the desired ideal braking-forcedistribution.

Axial displacement of the first cylinder end 12 to vary the value V₀ orl₀ is conveniently effected when the brake is not actuated, i.e., whenthe outlet pressure is P₂ =0. In this condition, the second piston 8assumes its initial position, shown in the drawings (X=0), in which theapertures of the equalizing channel 20, leading from the compressionchamber 11, is still open. The pressure connection then existing betweenthe compression chamber 11 and the chamber 30 makes possible areactionless displacement of the first cylinder end 12. It is, ofcourse, also possible to displace the first cylinder bottom 12 during abraking process (X=0). This condition, however, requires a higherexpenditure of force, so that it is then advisable to accomplish thechange in position of the first cylinder end 12 in a self lockingmanner.

Equations 1-4 have been set forth primarily to make clear that the useof the pneumatic elastic mechanism 11 having a progressivecharacteristic results in a continuous but nonlinear dependence betweenthe inlet pressure P₁ and the outlet pressure P₂. The equationsthemselves, however, apply only in first approximation, since, forexample, the effective areas of the control valve are not furtherconsidered. Practical considerations will show, however, thatoperability of the brake pressure regulator requires that the effectivearea F₃ of the second piston 8 (or in modified embodiments acorresponding area) must be greater than the effective valve area of thecontrol valve 7, so that the control valve may close after it has beenopened.

FIGS. 5 to 7 illustrate three representative embodiments of brakepressure regulators by which a family of characteristic curves eachfollowing a degressive course may be obtained. A preferred embodiment isshown in FIG. 7.

A degressive family of characteristic curves obtainable with the saidbrake pressure regulators is illustrated in FIG. 8. In FIG. 8, the solidline curve represents the characteristic line applying at "full load"condition resulting in positioning of the end member 12 at the "l₀ max"position and the dashed line curve corresponds to the "emptyload+driver" condition resulting in positioning of the end member 12 atthe "l₀ min" position. The intermediate dot-dash lines representcharacteristic curves corresponding to vehicle loads between full loadand empty load.

In the FIGS. 5-7 embodiments, the nonhydraulic first force mechanismhaving a linear characteristic, i.e., the mechanical elastic device 10"or 10"' acts--unlike the embodiments of FIGS. 2 and 3--not on the firstpiston 6', 6" or 6"' but, rather on the second piston 8', 8" or 8"' andthe nonhydraulic second force mechanism having a progressivecharacteristics, i.e., the pneumatic elastic mechanism 11', 11" or 11"'acts on the first piston 6', 6" or 6"' rather than on the second piston.Otherwise, however, the brake pressure regulators of FIGS. 5-7 areessentially identical, in their design and with reference to theirmethod of operation, to those of FIGS. 2 and 3. Accordingly, it will benecessary to describe only one of these embodiments in detail, likeelements in the three being provided with like primed, double primed ortriple primed reference numerals.

Thus, the difference between the FIG. 5 embodiment and the embodimentsof FIGS. 2 and 3 is that the nonhydraulic first force mechanism having alinear characteristic, i.e., the mechanical elastic mechanism 10', actson the second piston 8' and the nonhydraulic second force mechanismhaving a progressive characteristic, namely the pneumatic elasticmechanism with the compression chamber 11', acts on the first piston 6'.Here, too, the compression chamber 11' of the pneumatic elasticmechanism is designed as a cylindrical chamber with an axiallydisplaceable first cylinder end 12', the axial position of which isdisplaceable as a function of a vehicle parameter, in particular, as afunction of the load of the vehicle. In contrast to the embodiments ofFIGS. 2 and 3, the cylindrical chamber on the side facing the firstcylinder end 12' is limited axially not by the second piston 8' itself,but by an axially displaceable second cylinder end 13', which is rigidlyconnected to the first piston 6'. For this purpose it has a cylindricalextension 14' projecting away from the compression chamber 11', the freeend of which extension rests against the first piston 6' and forms aguide cylinder 15', within which the second piston 8' is axiallydisplaceable. In the cylindrical wall of the extension 14' there areopenings 16' which provide a constant connection between the secondpressure chamber 4' and the pressure outlet 5'.

In operation of the embodiment of FIGS. 5-7, when the brake pedal isactuated, the inlet pressure P₁, produced by the main brake cylinder ofthe motor vehicle and effective in the first pressure chamber 2 of thebrake pressure regulator, acts on the first piston 6', whereby thatpiston is deflected in X direction. The valve member 71' of the controlvalve 7' is thereby lifted from its valve seat so that the valve opening72' is exposed and the hydraulic fluid flows into the second pressurechamber 4', producing an outlet pressure P₂ in that chamber. Thedeflection of the first piston 6' in the X direction likewise causes thesecond cylinder end 13', which engages the piston 6 through thecylindrical extension 14', to be simultaneously displaced by the samedistance in the X direction. This causes the aperture of theequalization channel 20', opening into the compression chamber 11', tobe closed off. In accordance with the displacement of the secondcylinder end 13', the volume of the compression chamber 11' is reducedby the factor (l₀ -X), whereby the pressure in the compression chamberrises, according to the polytropic equation, from its initial pressureP₀ to: ##EQU4##

The pressure P_(i), rising progressively with the deflection of thesecond cylinder end 13', exerts a progressively rising opposing forceP_(i) ·F₄ on the second cylinder end, F₄ being the effective area of thesecond cylinder end. In contrast to the embodiments of FIGS. 2 and 3,the first piston 6' is thus not acted on by an elastically-inducedlinear opposing force, but, rather, by an elastically-induced opposingforce.

The valve member 71' in the first piston 6' is rigidly connected, by wayof the rod-shaped force-transmitting member 9', with the second piston8', which is within the guide cylinder 15'. Since the second piston 8',acted on by the force P₂ ·F₃, is likewise deflected in the X directionagainst the force of the mechanical elastic mechanism 10', the valvebore 72' is eventually shut off by the valve member 71'.

In the state of equilibrium the two pistons 6' and 8' are deflected thesame distance and the control valve 7' is closed. The following forceequations then apply to the two pistons 6' and 8':

    Piston 6': P.sub.1 ·F.sub.1 -P.sub.2 ·F.sub.2 -P=0, (5)

where approximating,

    P=P.sub.i ·F.sub.4 -P.sub.0 ·F.sub.4

applies.

Taking into account that, on a polytropic change of state in thecompression chamber 11', ##EQU5## it follows that: ##EQU6## The equation

    P.sub.2 ·F.sub.3 -k·X-P.sub.0 ·F.sub.3 =0 (7)

correspondingly applies to the second piston 8'. Since in the state ofequilibrium the two pistons 6' and 8' are deflected equidistantly, thedistance X in Equations 6 and 7 is the same. From this the equation forthe relation between inlet pressure P₁ and the outlet pressure P₂ may beobtained: ##EQU7##

Since the brake pressure regulator described in Equation 8 is technicalsignificant only when the denominator of the expression in parenthesisis positive, the value ##EQU8## represents the greatest possible outletpressure of the brake pressure regulator.

Equations 5 to 8 likewise apply only in first approximation, as somesimplifications have been made. Among other things, it has been assumedthat the initial pressure P₀ prevailing in the guide cylinder 15'retains its value corresponding to atmospheric pressure, even when thesecond piston 8' moves. If the mechanical elastic mechanism 10' has alarge force constant k, this assumption may be considered valid.However, if the force constant k is comparatively small and thereforethe product k·X is not in every case much greater than the product P_(i)·F₄, care must be taken to see that the guide cylinder 15' has aconnection to the outside, or else the characteristic of the regulatorwill change somewhat. If such an equalizing connection is not provided,as in FIGS. 5 and 6, then it is advantageous to arrange a check valve21' or 21' in the second cylinder end 13', or 13' by way of which anyexcess pressure with respect to the compression chamber 11' or 11" whenthe brake is not actuated may be eliminated.

Appropriate sizing of the various regulator parameters makes it possibleto establish the family of characteristic curves of the brake pressureregulator within a very wide range, the comparatively strongly curvedcourses of the characteristic lines in FIG. 8 being quasilinear in theboundary case. With the brake pressure regulator according to theinvention, therefore, it is possible to obtain, in simple fashion, afamily of characteristic curves coming very close to the idealbraking-force distribution of FIG. 1.

The embodiments shown in FIGS. 6 and 7 operate in a manner similar toFIG. 5 as described above.

In FIGS. 5-7, as in the embodiments of FIGS. 2 and 3, the size of thecompression chamber 11' 11" or 11"' is variable, in particular, as afunction of the vehicle load. The initial axial length l₀ of thischamber is readjusted by axial displacement of the first cylinder end12', 12" or 12"' which in the examples shown in FIGS. 5 and 6 isconnected by way of a coupling rod 23' or 23" with an auxiliary piston22' or 22" arranged within an auxiliary cylinder 31' or 31". Theauxiliary piston is displaced axially within the cylinder 31' or 31"under the influence of a pressure medium supplied by way of theadjusting inlet 19' or 19" against the effect of an elastic mechanism24' or 24". When the pressure in the auxiliary cylinder 31' or 31" ismade proportional to the vehicle load, the second cylinder end 12' or12" is in each instance displaced axially such that the magnitude of l₀is proportional to the corresponding vehicle load.

As is easily seen in FIGS. 2-6, the regulating process is alwaysassociated with an increase in the volume of the space receiving thehydraulic pressure fluid. This increase in volume amounts to

    ΔV=X·F.sub.3

and becomes small when either the displacement path X or the effectivearea F₃ of the second piston 8, 8' or 8" is kept small. Generally, onthe one hand, the smaller the effective area F₃ and the greater theforce constant k of the mechanical elastic mechanism 10, 10' or 10" ismade, the smaller the displacement path X becomes; and, generally thegreater the force constant k and the smaller the effective area F₃, thegreater the highest attainable outlet pressure P₂ becomes. It istherefore advantageous to make the force constant k large and theeffective area F₃ small, so that the brake pressure regulator can bemade small enough that the "hydraulic dead space" X·F₃, is insignificantwith respect to any elasticities present in the brake circuit.

A brake pressure regulator modified in this way is represented in theembodiment of FIG. 7. In that brake pressure regulator the insidediameter of the guide cylinder 15"', in the region in which the secondpiston 8"' is located, is made substantially smaller than the diameterof the first pressure chamber 2 accomodating the first piston 6"'. The"hydraulic dead space" ΔV=X·F₃ is thus comparatively small in thisembodiment.

When, in an optimum brake pressure regulator design, the displacementpath X becomes so small that the aperture of the equalizing channel 20"opening into the compression chamber 11" can no longer be closedcompletely, then it is advantageous to provide, within the equalizingchannel 20", a valve such as an electromagnetic valve or the like asindicated by the valve 25" in the embodiment of FIG. 6. The valve may beactuated by, for example, closing of the usual brake-light switch or,alternatively, when a selected, comparatively small, pressure in thebrake circuit is exceeded. For the sake of completeness, additionalequalizing channels 27"' are indicated in the embodiment of FIG. 7,through which the guide cylinder 15"' communicates with the atmosphere.As the result, a special check valves like 21' and 21" in the secondcylinder ends 13' and 13" of the embodiments shown in FIGS. 5 and 6 arenot necessary. In addition, a check valve 26"' is provided between thepressure inlet 3"' and the pressure outlet 5, to assure that thepressure at the pressure outlet 5 can never become greater than that atthe pressure inlet 3"'.

As is shown in the embodiment of FIG. 6, it is alternatively possible,in principle, to omit all sealing rings when the pistons, piston rodsand cylindrical end members, etc. to be sealed are suspended in axiallyelastic diaphragms 17". The operation of this brake pressure regulatoris the same as that of the regulator represented in FIG. 5. In thisembodiment the electromagnetic valve 25" as shown in the drawing must beprovided since the channel 20" can not otherwise be closed.

It is also possible by any conventional arrangement, not shown in thedrawings, to prevent the basic or initial pressure P₀ of the compressionchamber 11, 11', 11" or 11"' from dropping below a given value, so thatthe regulating characteristic of the brake pressure regulator may bevery strong.

I claim:
 1. A hydraulic brake pressure regulator for the hydraulic brakesystem of a motor vehicle providing a family of characteristic curveswhich are dependent upon a vehicle parameter wherein each of thecharacteristic curves follows a nonlinear and continuous coursecomprising a housing having a first pressure chamber with a pressureinlet and a second pressure chamber with a pressure outlet, the firstand second chambers being separated from one another by an axiallydisplaceable first piston having a control valve which permits apressure and flow connection for a hydraulic pressure medium between thefirst and second pressure chambers to be opened and closed, a secondpiston which is coaxial with respect to said first piston and axiallydisplaceable independently of the first piston and which defines thesecond pressure chamber in the axial direction, the second piston beingcoupled by way of a rigid, force-transmitting member with the controlvalve so that opening and closing of the control valve is dependent uponthe axial separation of the two pistons, a non-hydraulic first forcemechanism having a linear characteristic, a non-hydraulic second forcemechanism independent of the first force mechanism and having avariable, progressive characteristic dependent on a parameter of thevehicle, one of the force mechanisms acting on the first piston in adirection opposing the inlet pressure and the other force mechanismacting on the second piston in the same effective direction.
 2. Ahydraulic brake pressure regulator according to claim 1 for insertionbetween a main brake cylinder and a wheel brake cylinder of the rearaxle of the vehicle wherein the nonhydraulic first force mechanism actson the second piston and the nonhydraulic second force mechanism acts onthe first piston.
 3. A hydraulic brake pressure regulator according toclaim 1 for insertion between a main brake cylinder and a wheel brakecylinder of the front axle of the vehicle wherein the nonhydraulic firstforce mechanism acts on the first piston and the nonhydraulic secondforce mechanism acts on the second piston.
 4. A hydraulic brake pressureregulator according to claim 1 wherein the nonhydraulic second forcemechanism comprises a pneumatic elastic mechanism including acompression chamber filled with a compressible gaseous medium having asize which is variable in accordance with a vehicle parameter.
 5. Ahydraulic brake pressure regulator according to claim 4 wherein the sizeof the compression chamber of the pneumatic elastic mechanism isvariable in accordance with the vehicle load so that with no load it hasa minimum value and with full load it has a maximum value.
 6. Ahydraulic brake pressure regulator according to claim 4 wherein thecompression chamber comprises a cylindrical chamber having an axiallydisplaceable first cylinder end, the position of which is dependent on avehicle parameter, and an axially displaceable second cylinder end,which has a rigid connection with one of the two pistons.
 7. A hydraulicbrake pressure regulator according to claim 6 wherein the axiallydisplaceable cylinder end has a cylindrical extension projecting awayfrom the compression chamber forming a guide cylinder for the secondpiston, the end of the cylindrical extension engaging the first pistonand an opening being provided in the wall of the cylindrical extensionwhich provides a constant connection between the second pressure chamberand the pressure outlet.
 8. A hydraulic brake pressure regulatoraccording to claim 7, wherein the inside diameter of the guide cylinderat least in the region containing the second piston is substantiallysmaller than the diameter of the first pressure chamber which receivesthe first piston.
 9. A hydraulic brake pressure regulator according toclaim 6, wherein the axially displaceable end of the second cylinder isformed by the second piston.
 10. A hydraulic brake pressure regulatoraccording to claim 1 wherein at least some of said first and secondpistons and said first and second axially displacecylinder ends aresuspended in axially elastic diaphragms which, in addition to providingsupport, also provide a seal and separate the regions on the oppositesides thereof.
 11. A hydraulic brake pressure regulator for thehydraulic brake system of a motor vehicle providing a family ofcharacteristic curves which are dependent upon a vehicle parameterwherein each of the characteristic curves follows a nonlinear andcontinuous course comprising a housing having a first pressure chamberwith a pressure inlet and a second pressure chamber with a pressureoutlet, the first and second chambers being separated from one anotherby an axially displaceable first piston having a control valve whichpermits a pressure and flow connection for a hydraulic pressure mediumbetween the first and second pressure chambers to be opened and closed,a second piston which is coaxial with respect to said first piston andaxially displaceable independently of the first piston and which definesthe second pressure chamber in the axial direction, the second pistonbeing coupled by way of a rigid, force-transmitting member with thecontrol valve so that opening and closing of the control valve isdependent upon the axial separation of the two pistons, a non-hydraulicfirst force mechanism having a linear characteristic and acting on saidfirst piston in a direction opposing the inlet pressure, a non-hydraulicsecond force mechanism independent of the first force mechanism andhaving a variable, progressive characteristic dependent on a parameterof the vehicle acting on said second piston in the same effectivedirection as said first piston, said second force mechanism including acylindrical compression chamber filled with a compressible gas andhaving an axially displaceable first cylinder end, the position of whichis dependent on a vehicle parameter, and an axially displaceable secondcylinder end which has a rigid connection with one of said two pistons,said axially displaceable second cylinder end having a cylindricalextension projecting away from the compression chamber and forming, forthe second piston, a guide cylinder of substantially smallercross-section than the cross-section of the first pressure chamber inwhich the first piston is disposed, the end of said cylindricalextension engaging the first piston and an opening being provided in thewall of the cylindrical extension which provides a constant connectionbetween the second pressure chamber and said pressure outlet.
 12. Ahydraulic brake pressure regulator according to claim 11 wherein saidfirst cylinder end is coupled to a third piston axially displaceablewithin an auxiliary cylinder against the effect of an elastic medium inresponse to a pressure medium proportional to vehicle load.
 13. Ahydraulic brake pressure regulator according to claim 12 wherein atleast some of said first, second and third pistons and first and secondcylinder ends are suspended in axially elastic diaphragms providingsupport and a seal therefor and separating the regions on opposite sidesthereof, an equalizing channel is provided communicating with thecompression chamber, with valve means responsive to actuation of thevehicle brake system for closing said equalizing channel.